Double-wall-tube heat exchanger

ABSTRACT

A double-wall-tube heat exchanger has an outer tube, and an inner tube disposed concentrically in and spaced apart from the outer tube. A clearance between the outer tube and the inner tube and the interior of the inner tube serve as respective refrigerant flow paths. The inner tube has a plurality of interior fins formed on its inner circumferential surface. The interior fins project radially inward; extend in a longitudinal direction of the inner tube; and are arranged at circumferential intervals. The inner tube has a plurality of elongated projections formed on its outer circumferential surface. The elongated projections project radially outward; extend in the longitudinal direction; and are arranged at circumferential intervals. The interior fins are greater in height than the elongated projections. The double-wall-tube heat exchanger exhibits excellent heat exchange performance.

BACKGROUND OF THE INVENTION

The present invention relates to a double-wall-tube heat exchanger and, more particularly, to a double-wall-tube heat exchanger which has an outer tube, and an inner tube provided in and spaced apart from the outer tube.

Herein, the term “condenser” refers to not only an ordinary condenser but also a subcooling condenser, which has a condensing section and a subcooling section.

A conventionally proposed refrigeration system for use in a car air conditioner includes a compressor; a condenser having a condensing section and a subcooling section; an evaporator; an expansion valve serving as a pressure-reducing device; a vapor-liquid separator; and an intermediate heat exchanger disposed between the condenser and the evaporator and adapted to perform heat exchange between a high-temperature refrigerant from the subcooling section of the condenser and a low-temperature refrigerant from the evaporator (as disclosed in, for example, Japanese Patent Application Laid-Open (kokai) No. 2006-162241). In the refrigeration system described in the publication, the refrigerant which has been subcooled in the subcooling section of the condenser is further cooled in the intermediate heat exchanger by the low-temperature, low-pressure refrigerant from the evaporator. By this procedure, the cooling performance of the evaporator is improved.

The intermediate heat exchanger used in the refrigeration system described in the above-mentioned publication has an outer tube, and an inner tube disposed in and spaced apart from the outer tube; the inner tube has grooves which are formed on its outer wall surface, by deforming its wall, in such a manner as to extend in its longitudinal direction; a clearance between the outer tube and the inner tube serves as a high-temperature refrigerant flow path through which the high-temperature refrigerant from the condenser flows; and the interior of the inner tube serves as a low-temperature refrigerant flow path through which the low-temperature refrigerant from the evaporator flows.

However, the intermediate heat exchanger described in the above-mentioned publication involves the following problem: the area of heat transfer between the high-temperature refrigerant flow path and the low-temperature refrigerant flow path is small, resulting in insufficient heat exchange performance.

SUMMARY OF THE INVENTION

An object of the present invention is to solve the above-mentioned problem and to provide a double-wall-tube heat exchanger exhibiting excellent heat exchange performance.

To achieve the above object, the present invention comprises the following modes.

1) A double-wall-tube heat exchanger comprising an outer tube, and an inner tube disposed in and spaced apart from the outer tube. A clearance between the outer tube and the inner tube and the interior of the inner tube serve as respective refrigerant flow paths. The inner tube has a plurality of interior fins formed on the inner circumferential surface thereof. The interior fins project radially inward; extend in the longitudinal direction of the inner tube; and are arranged at circumferential intervals. The inner tube has a plurality of elongated projections formed on the outer circumferential surface thereof. The elongated projections project radially outward; extend in the longitudinal direction; and are arranged at circumferential intervals. The fin height of the interior fins is greater than the projecting height of the elongated projections.

2) A double-wall-tube heat exchanger according to par. 1), wherein a radial clearance between the inner circumferential surface of the outer tube and a portion of the outer circumferential surface of the inner tube where the elongated projections are not formed is 0.4 mm to 1.2 mm inclusive.

Par. 2) specifies a radial clearance of 0.4 mm to 1.2 mm inclusive between the inner circumferential surface of the outer tube and the portion of the outer circumferential surface of the inner tube where the elongated projections are not formed, for the following reason. If the radial clearance is excessively small, pressure loss increases sharply in the refrigerant flow path formed between the inner tube and the outer tube. On the other hand, if the radial clearance is excessively large, the flow velocity of a refrigerant drops in the refrigerant flow path formed between the inner tube and the outer tube, potentially resulting in a drop in heat transfer coefficient.

3) A double-wall-tube heat exchanger according to par. 1), wherein a clearance between projecting ends of the elongated projections of the inner tube and the inner circumferential surface of the outer tube is 0 mm to 0.5 mm inclusive.

Par. 3) specifies a clearance of 0 mm to 0.5 mm inclusive between the projecting ends of the elongated projections of the inner tube and the inner circumferential surface of the outer tube, for the following reason. If the clearance is excessively large, in the case where the double-wall-tube heat exchanger has a bend(s), wrinkles are apt to be formed on the outer tube in a bending process.

4) A double-wall-tube heat exchanger comprising an outer tube, and an inner tube disposed in and spaced apart from the outer tube. A clearance between the outer tube and the inner tube and the interior of the inner tube serves as respective refrigerant flow paths. The inner tube has a plurality of interior fins formed on the inner circumferential surface thereof. The interior fins project radially inward; extend in the longitudinal direction of the inner tube; and are arranged at circumferential intervals. The outer tube has a plurality of elongated projections formed on the inner circumferential surface thereof. The elongated projections project radially inward; extend in the longitudinal direction of the outer tube; and are arranged at circumferential intervals.

5) A double-wall-tube heat exchanger according to par. 4), wherein a radial clearance between a portion of the inner circumferential surface of the outer tube where the elongated projections are not formed and the outer circumferential surface of the inner tube is 0.4 mm to 1.2 mm inclusive.

Par. 5) specifies a radial clearance of 0.4 mm to 1.2 mm inclusive between the portion of the inner circumferential surface of the outer tube where the elongated projections are not formed and the outer circumferential surface of the inner tube, for the following reason. If the radial clearance is excessively small, pressure loss increases sharply in the refrigerant flow path formed between the inner tube and the outer tube. On the other hand, if the radial clearance is excessively large, the flow velocity of a refrigerant drops in the refrigerant flow path formed between the inner tube and the outer tube, potentially resulting in a drop in heat transfer coefficient.

6) A double-wall-tube heat exchanger according to par. 4), wherein a clearance between projecting ends of the elongated projections of the outer tube and the outer circumferential surface of the inner tube is 0 mm to 0.5 mm inclusive.

Par. 6) specifies a clearance of 0 mm to 0.5 mm inclusive between the projecting ends of the elongated projections of the outer tube and the outer circumferential surface of the inner tube, for the following reason. If the clearance is excessively large, in the case where the double-wall-tube heat exchanger has a bend(s), wrinkles are apt to be formed on the outer tube in a bending process.

7) A double-wall-tube heat exchanger according to par. 1) or 4), wherein the interior fins of the inner tube have a fin thickness of 0.2 mm to 2.0 mm inclusive.

Par. 7) specifies a fin thickness of 0.2 mm to 2.0 mm inclusive for the interior fins of the inner tube, for the following reason. If the fin thickness is excessively thin, the fin efficiency of the interior fins drops, and working may become difficult. If the fin thickness is in excess of 2.0 mm, the effect of improving the fin efficiency of the interior fins is impaired, and working may become difficult. In consideration of extrusion workability in forming the inner tube by extrusion, and bending workability in the case where the double-wall-tube heat exchanger has a bend(s), the fin thickness of the interior fins of the inner tube is more preferably 0.3 mm to 0.7 mm inclusive.

8) A double-wall-tube heat exchanger according to par. 1) or 4), wherein the interior fins have a fin height of 1.0 mm to 3.0 mm inclusive.

Par. 8) specifies a fin height of 1.0 mm to 3.0 mm inclusive for the interior fins of the inner tube, for the following reason. If the fin height is excessively low, the area of heat transfer between the inner tube and a refrigerant which flows through the refrigerant flow path in the inner tube fails to become sufficiently large; as a result, heat transfer performance fails to be sufficiently improved. If the fin height is excessively high, in the case where the double-wall-tube heat exchanger has a bend(s), the interior fins may be buckled in a bending process, potentially blocking the refrigerant flow path in the inner tube.

9) A double-wall-tube heat exchanger according to par. 1) or 4), wherein the inner tube has an inside diameter of 12 mm or greater.

Par. 9) specifies an inside diameter of 12 mm or greater for the inner tube, for the following reason. If the inside diameter of the inner tube is excessively small, pressure loss in the refrigerant flow path in the inner tube increases sharply.

In the double-wall-tube heat exchanger according to any one of pars. 1) to 9), the interior fins of the inner tube may be arranged at a fin pitch of 2 mm or greater as measured at roots of the interior fins, for the following reason. If the fin pitch is excessively small, pressure loss in the refrigerant flow path in the inner tube increases sharply. Particularly, in the case where the double-wall-tube heat exchanger has a bend(s), the interior fins come into contact with one another in a bending process, causing a sharp increase in pressure loss in the refrigerant flow path in the inner tube.

In the double-wall-tube heat exchanger according to any one of pars. 1) to 9), the inner tube may have a wall thickness of 0.2 mm to 2.0 mm inclusive, for the following reason. If the wall thickness is excessively thin, strength becomes insufficient. If the wall thickness is excessively thick, weight increases, and cost increases as well.

According to the double-wall-tube heat exchanger of par. 1), the inner tube has a plurality of the interior fins formed on the inner circumferential surface thereof, the interior fins projecting radially inward, extending in the longitudinal direction of the inner tube, and being arranged at circumferential intervals; the inner tube has a plurality of the elongated projections formed on the outer circumferential surface thereof, the elongated projections projecting radially outward, extending in the longitudinal direction, and being arranged at circumferential intervals; and the fin height of the interior fins is greater than the projecting height of the elongated projections. Thus, as compared with the double-wall-tube heat exchanger described in the above-mentioned publication, the area of heat transfer between the refrigerant flow path formed between the inner and outer tubes and the refrigerant flow path in the inner tube is greater, so that heat exchange performance is improved. In the case where the double-wall-tube heat exchanger of par. 1) is used as an intermediate heat exchanger of a refrigeration system described in the above-mentioned publication, a vapor-phase refrigerant, whose heat transfer coefficient is relatively low, flows through the refrigerant flow path in the inner tube. Since the interior fins function to increase the area of heat transfer in the refrigerant flow path in the inner tube through which the vapor-phase refrigerant flows, the performance of the double-wall-tube heat exchanger is improved. In the case where the double-wall-tube heat exchanger has a bend(s), the elongated projections function to prevent crushing of the refrigerant flow path formed between the inner tube and the outer tube.

According to the double-wall-tube heat exchanger of par. 2), an increase in pressure loss in the refrigerant flow path formed between the inner and outer tubes can be prevented, and the flow velocity of a refrigerant in the refrigerant flow path formed between the inner and outer tubes increases, whereby heat transfer coefficient is improved, resulting in improvement of the performance of the double-wall-tube heat exchanger.

According to the double-wall-tube heat exchanger of par. 3), in the case where the double-wall-tube heat exchanger has a bend(s), formation of wrinkles on the outer tube in a bending process can be reliably prevented.

According to the double-wall-tube heat exchanger of par. 4), the inner tube has a plurality of the interior fins formed on the inner circumferential surface thereof, the interior fins projecting radially inward, extending in the longitudinal direction of the inner tube, and being arranged at circumferential intervals; and the outer tube has a plurality of the elongated projections formed on the inner circumferential surface thereof, the elongated projections projecting radially inward, extending in the longitudinal direction of the outer tube, and being arranged at circumferential intervals. Thus, as compared with the double-wall-tube heat exchanger described in the above-mentioned publication, the area of heat transfer between the refrigerant flow path formed between the inner and outer tubes and the refrigerant flow path in the inner tube is greater, so that heat exchange performance is improved. Also, in the case where the double-wall-tube heat exchanger has a bend(s), the elongated projections function to prevent crushing of the refrigerant flow path formed between the inner tube and the outer tube. Furthermore, since the outer surface of the inner tube assumes the form of a smooth cylindrical surface, joining work can be facilitated for a refrigerant inflow pipe, a refrigerant outflow pipe, a joint member, etc.

According to the double-wall-tube heat exchanger of par. 5), an increase in pressure loss in the refrigerant flow path formed between the inner and outer tubes can be prevented, and the flow velocity of a refrigerant in the refrigerant flow path formed between the inner and outer tubes increases, whereby heat transfer coefficient is improved, resulting in improvement of the performance of the double-wall-tube heat exchanger.

According to the double-wall-tube heat exchanger of par. 6), in the case where the double-wall-tube heat exchanger has a bend(s), formation of wrinkles on the outer tube in a bending process can be reliably prevented.

According to the double-wall-tube heat exchanger of par. 7), the fin efficiency of the interior fins of the inner tube is improved, whereby heat exchange performance is improved.

According to the double-wall-tube heat exchanger of par. 8), the area of heat transfer between the inner tube and a refrigerant which flows through the refrigerant flow path in the inner tube becomes sufficiently large, whereby the performance of heat transfer between the refrigerant and the inner tube is sufficiently improved. In the case where the double-wall-tube heat exchanger has a bend(s), there is prevented blocking of the refrigerant flow path in the inner tube which could otherwise result from the interior fins being buckled in a bending process.

According to the double-wall-tube heat exchanger of par. 9), an increase in pressure loss in the refrigerant flow path in the inner tube can be restrained.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a partially cutaway front view showing the configuration of a double-wall-tube heat exchanger according to Embodiment 1 of the present invention with a longitudinally intermediate portion omitted;

FIG. 2 is an enlarged fragmentary view of FIG. 1;

FIG. 3 is a sectional view taken along line A-A of FIG. 2;

FIG. 4 is an enlarged fragmentary view of FIG. 3;

FIG. 5 is a sectional view taken along line BPB of FIG. 2;

FIG. 6 is an enlarged fragmentary view of FIG. 5;

FIG. 7 is a diagram showing a refrigeration system which uses the double-wall-tube heat exchanger of Embodiment 1 as an intermediate heat exchanger;

FIG. 8 is a graph showing the relation between the fin thickness and the fin efficiency of the interior fins;

FIG. 9 is a graph showing the relation of the number of fins and the fin pitch to exchanged heat quantity and pressure loss;

FIG. 10 is a graph showing the relation of the inside diameter of the inner tube to pressure loss and overall heat transfer coefficient;

FIG. 11 is a graph showing the relation of liquid flow path width to pressure loss and overall heat transfer coefficient;

FIG. 12 is a perspective, fragmentary view showing a modified inner tube of the double-wall-tube heat exchanger of Embodiment 1;

FIG. 13 is an equivalent view of FIG. 2, showing a double-wall-tube heat exchanger according to Embodiment 2 of the present invention;

FIG. 14 is a sectional view taken along line C-C of FIG. 13;

FIG. 15 is an enlarged fragmentary view of FIG. 14;

FIG. 16 is a sectional view taken along line D-D of FIG. 13;

FIG. 17 is an enlarged fragmentary view of FIG. 16; and

FIG. 18 is an equivalent view of FIG. 13, showing a modified inner tube of the double-wall-tube heat exchanger of Embodiment 2.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

Embodiments of the present invention will next be described in detail with reference to the drawings.

In the following description, the term “aluminum” encompasses aluminum alloys in addition to pure aluminum.

In the drawings, like sections or components throughout the several views are denoted by like reference numerals, and repeated description thereof is omitted.

Embodiment 1

The present embodiment is shown in FIGS. 1 to 7.

FIG. 1 shows the configuration of a double-wall-tube heat exchanger according to Embodiment 1 of the present invention. FIGS. 2 to 6 show the configurations of essential portions of the double-wall-tube heat exchanger. FIG. 7 shows a refrigeration system which uses the double-wall-tube heat exchanger of Embodiment 1 as an intermediate heat exchanger.

In FIGS. 1 to 5, a double-wall-tube heat exchanger 1 includes an outer tube 2 and an inner tube 3. The outer tube 2 is formed of an aluminum extrudate having a circular cross section. The inner tube 3 is formed of an aluminum extrudate having a circular cross section and is inserted concentrically into and spaced apart from the outer tube 2. A clearance between the outer tube 2 and the inner tube 3 serves as a first refrigerant flow path 4. The interior of the inner tube 3 serves as a second refrigerant flow path 5.

The outer tube 2 has expanded tube portions 6 and 7 located slightly longitudinally inward of the opposite ends thereof. A refrigerant inlet (not shown) is formed in the tube wall of the expanded tube portion 6 of the outer tube 2. A refrigerant outlet 8 is formed in the tube wall of the expanded tube portion 7 of the outer tube 2. An end portion of a liquid-phase refrigerant inflow pipe 9 made of aluminum is inserted into the refrigerant inlet and is brazed to the expanded tube portion 6. An end portion of a liquid-phase refrigerant outflow pipe 11 made of aluminum is inserted into the refrigerant outlet 8 and is brazed to the expanded tube portion 7. Preferably, the outer tube 2 has an outside diameter of 25 mm or less and a tube wall thickness of 0.2 mm to 2.0 mm inclusive.

The inner tube 3 has a plurality of interior fins 12 formed integrally with the inner circumferential surface thereof. The interior fins 12 project radially inward, extend in the longitudinal direction of the inner tube 3, and are arranged at equal circumferential intervals. The inner tube 3 also has a plurality of elongated projections 13 formed integrally with the outer circumferential surface thereof. The elongated projections 13 project radially outward, extend in the longitudinal direction, and are arranged at equal circumferential intervals. The fin height of the interior fins 12 is greater than the projecting height of the elongated projections 13. Portions of the outer tube 2 which are located longitudinally outward of the expanded portions 6 and 7 are subjected to roller working which is performed from the radial outside toward the radial inside along the entire circumference, thereby forming diameter-reduced portions 14. The diameter-reduced portions 14 are brazed to the inner tube 3 at positions located toward the opposite ends of the inner tube 3. The diameter-reduced portions 14 are formed after the inner tube 3 is placed within the outer tube 2. In the course of formation of the diameter-reduced portions 14, associated portions of the elongated projections 13 of the inner tube 3 are crushed and bite into the inner circumferential surfaces of the diameter-reduced portions 14 (see FIGS. 5 and 6). This reduces the clearance between the inner circumferential surface of the outer tube 2 and a portion of the outer circumferential surface of the inner tube 3 where the elongated projections 13 are not formed, to such an extent as to be filled with a brazing material. In this condition, the diameter-reduced portions 14 of the outer tube 2 are brazed to the inner tube 3. Also, the clearance between the inner circumferential surfaces of the diameter-reduced portions 14 of the outer tube 2 and the portion of the outer circumferential surface of the inner tube 3 where the elongated projections 13 are not formed is filled with a brazing material 17 (see FIG. 6).

An end portion; i.e., an expanded pipe portion 15 a, of a vapor-phase refrigerant inflow pipe 15 made of aluminum is fitted onto and brazed to an end portion of the inner tube 3 which is located on a side toward the refrigerant outlet 8. Similarly, an end portion; i.e., an expanded pipe portion 16 a, of a vapor-phase refrigerant outflow pipe 16 made of aluminum is fitted onto and brazed to an end portion of the inner tube 3 which is located on a side toward the refrigerant inlet. The portions of the inner tube 3 onto which the expanded pipe portions 15 a and 16 a are fitted have associated portions of the elongated projections 13 cut off. Instead of cutting off the associated portions of the elongated projections 13, as in the case of brazing between the outer tube 2 and the inner tube 3, the expanded pipe portion 15 a of the vapor-phase refrigerant inflow pipe 15 and the expanded pipe portion 16 a of the vapor-phase refrigerant outflow pipe 16 may be pressed from the radial outside so as to establish a condition in which associated portions of the elongated projections 13 are crushed and bite into the inner circumferential surfaces of the expanded pipe portions 15 a and 16 a; as a result, the clearance between the inner circumferential surfaces of the expanded pipe portions 15 a and 16 a and a portion of the outer circumferential surface of the inner tube 3 where the elongated projections 13 are not formed is reduced to such an extent as to be filled with the brazing material. Preferably, brazing between the inner tube 3 and each of the expanded pipe portions 15 a and 16 a of the vapor-phase refrigerant inflow and outflow pipes 15 and 16 is carried out simultaneously with brazing between the outer tube 2 and the inner tube 3 while an appropriate gap is maintained between the opposite ends of the outer tube 2 and the corresponding ends of the expanded pipe portions 15 a and 16 a of the vapor-phase refrigerant inflow and outflow pipes 15 and 16.

FIG. 7 shows a refrigeration system in which the above-described double-wall-tube heat exchanger 1 is used as an intermediate heat exchanger.

In FIG. 7, the refrigeration system uses, for example, a chlorofluorocarbon-based refrigerant. The refrigeration system includes a compressor 20; a condenser 21 having a condensing section 22, a liquid receiver 23 serving as a vapor-liquid separator, and a subcooling section 24; an evaporator 25; an expansion valve 26 serving as a pressure-reducing device; and the double-wall-tube heat exchanger 1 which serves as an intermediate heat exchanger for performing heat exchange between a refrigerant from the condenser 20 and a refrigerant from the evaporator 25. Piping extending from the subcooling section 24 of the condenser 20 is connected to the liquid-phase refrigerant inflow pipe 9 connected to the outer tube 2 of the double-wall-tube heat exchanger 1. Similarly, piping extending to the expansion valve 26 is connected to the liquid-phase refrigerant outflow pipe 11 connected to the outer tube 2. Also, piping extending from the evaporator 25 is connected to the vapor-phase refrigerant inflow pipe 15 connected to the inner tube 3 of the double-wall-tube heat exchanger 1. Similarly, piping extending to the compressor 20 is connected to the vapor-phase refrigerant outflow pipe 16 connected to the inner pipe 3. The refrigeration system is mounted in a vehicle; for example, an automobile, as a car air conditioner.

In operation of the refrigeration system, a high-temperature, high-pressure vapor-liquid mixed-phase refrigerant, which has undergone compression in the compressor 20, is cooled and condensed in the condensing section 22 of the condenser 21. Subsequently, the refrigerant flows into the liquid receiver 23 and is separated into two phases; namely, the vapor phase and the liquid phase. The resultant liquid-phase refrigerant flows into the subcooling section 24 and is subcooled. The subcooled liquid-phase refrigerant flows through the liquid-phase refrigerant inflow pipe 9 and flows into the first refrigerant flow path 4 of the double-wall-tube heat exchanger 1. At this time, by the effect of the expanded tube portion 6, the liquid-phase refrigerant dividedly flows into all the channels formed between the adjacent elongated projections 13 in the first refrigerant flow path 4. Meanwhile, a vapor-phase refrigerant from the evaporator 25 passes through the vapor-phase refrigerant inflow pipe 15 and flows into the second refrigerant flow path 5 of the double-wall-tube heat exchanger 1. While flowing through the first refrigerant flow path 4, the liquid-phase refrigerant is further cooled by the vapor-phase refrigerant whose temperature is relatively low and which flows through the second refrigerant flow path 5. Having passed through all the channels formed between the adjacent elongated projections 13 in the first refrigerant flow path 4 of the double-wall-tube heat exchanger 1, flows of the liquid-phase refrigerant join together in the expanded tube portion 7. The resultant liquid-phase refrigerant flows to the expansion valve 26 through the liquid-phase refrigerant outflow pipe 11. In the expansion valve 26, the liquid-phase refrigerant is adiabatically expanded and is thereby pressure-reduced. Subsequently, the two-phase refrigerant flows into the evaporator 25 and is evaporated in the evaporator 25. Meanwhile, having passed through the second refrigerant flow path 5 of the double-wall-tube heat exchanger 1, the vapor-phase refrigerant flows to the compressor 20 through the vapor-phase refrigerant outflow pipe 16.

In the double-wall-tube heat exchanger 1 in which the outer tube 2 and the inner tube 3 are formed of aluminum, the fin thickness T1 of the interior fins 12 of the inner tube 3 is preferably 0.2 mm to 2.0 mm inclusive. This has been obtained from the results of computer simulation calculation. Specifically, the relation between the fin thickness T1 and fin efficiency of the interior fins 12 was obtained by computer simulation calculation which was conducted while the fin thickness T1 and the fin height H1 of the interior fins 12 were varied under the following condition: the heat transfer coefficient in heat transfer from the vapor-phase refrigerant flowing through the second refrigerant flow path 5 to the inner circumferential surface of the inner tube 3 in the double-wall-tube heat exchanger 1 was set to 380 W/m²·K. FIG. 8 shows the results of the computer simulation calculation. It has been derived from the results shown in FIG. 8 that a fin thickness T1 of the interior fins 12 of 0.2 mm to 2.0 mm inclusive is preferred. As is apparent from FIG. 8, when the fin thickness T1 of the interior fins 12 is less than 0.2 mm, fin efficiency drops sharply. Also, when the fin thickness T1 is in excess of 1.2 mm, the effect of improving fin efficiency is saturated.

Preferably, the fin height H1 of the interior fins 12 of the inner tube 3 is 1.0 mm to 3.0 mm inclusive. This also has been derived from the results shown in FIG. 8.

Preferably, the fin pitch P1 of the interior fins 12 of the inner tube 3 as measured at roots of the interior fins 12 is 2 mm or greater. This has been obtained from the results of computer simulation calculation. Specifically, the relation of the number of the interior fins 12 to pressure loss and the exchanged heat quantity between the liquid-phase refrigerant and the vapor-phase refrigerant was obtained by computer simulation calculation which was conducted while the number of the interior fins 12 (the fin pitch P1 as measured at roots of the interior fins 12) was varied under the following conditions: the inside diameter D of the inner tube 3 was set to 13.5 mm; the fin thickness T1 of the interior fins 12 was set to 0.5 mm; the fin height H1 of the interior fins 12 was set to 1.5 mm; the elongated-projection thickness T2 of the elongated projections 13 was set to 0.5 mm; the projecting height H2 of the elongated projections 13 was set to 0.5 mm; the pitch P2 of the elongated projections 13 as measured at roots of the elongated projections 13 was set to 3.0 mm; the radial clearance W (hereinafter referred to as liquid flow path width) between the inner circumferential surface of the outer tube 2 and a portion of the outer circumferential surface of the inner tube 3 where the elongated projections 13 are not formed was set to 0.8 mm; the temperature and pressure of the liquid-phase refrigerant as measured at the inlet of the first refrigerant flow path 4 were set to 42.0° C. and 1.28 MPaG, respectively; and the temperature and pressure of the vapor-phase refrigerant as measured at the inlet of the second refrigerant flow path 5 were set to 8.0° C. and 0.21 MPaG, respectively. FIG. 9 shows the results of the computer simulation calculation. It has been derived from the results shown in FIG. 9 that a fin pitch P1 of the interior fins 12 as measured at roots of the interior fins 12 of 2.0 mm or greater is preferred. As is apparent from FIG. 9, when the fin pitch P1 of the interior fins 12 is less than 2 mm, pressure loss in the second refrigerant flow path 5 increases. In FIG. 9, exchanged heat quantity and pressure loss are expressed in percentage with those of the case where an inner tube not having the interior fins is used, being each taken as 100%.

Preferably, the inside diameter D of the inner tube 3 is 12 mm or greater. This has been obtained from the results of computer simulation calculation. Specifically, the relation of the inside diameter D of the inner tube 3 to pressure loss and the overall heat transfer coefficient in heat transfer from the vapor-phase refrigerant flowing through the second refrigerant flow path 5 to the inner circumferential surface of the inner tube 3 was obtained by computer simulation calculation which was conducted while the inside diameter D of the inner tube 3 was varied under the following conditions: the fin thickness T1 of the interior fins 12 of the inner tube 3 was set to 0.5 mm; the fin height H1 of the interior fins 12 was set to 1.5 mm; the fin pitch P1 of the interior fins 12 as measured at roots of the interior fins 12 was set to 2.5 mm; and the temperature and pressure of the vapor-phase refrigerant as measured at the inlet of the second refrigerant flow path 5 were set to 5.0° C. and 0.30 MPaG, respectively. FIG. 10 shows the results of the computer simulation calculation. It has been derived from the results shown in FIG. 10 that an inside diameter D of the inner tube 3 of 12 mm or greater is preferred. As is apparent from FIG. 10, when the inside diameter D of the inner tube 3 is less than 12 mm, pressure loss increases sharply. Preferably, the upper limit of the inside diameter D of the inner tube 3 is 18 mm. This is because, as is apparent from FIG. 10, when the inside diameter D of the inner tube 3 is in excess of 18 mm, the overall heat transfer coefficient in heat transfer from the vapor-phase refrigerant flowing through the second refrigerant flow path 5 to the inner circumferential surface of the inner tube 3 drops. In FIG. 10, pressure loss and the overall heat transfer coefficient in heat transfer from the vapor-phase refrigerant flowing through the second refrigerant flow path 5 to the inner circumferential surface of the inner tube 3 are expressed in percentage with those of the case where an inner tube having an inside diameter of 12 mm is used, being each taken as 100%.

Further, preferably, the liquid flow path width W is 0.4 mm to 1.2 mm inclusive. This has been obtained from the results of computer simulation calculation. Specifically, the relation of the liquid flow path width W to pressure loss and the overall heat transfer coefficient in heat transfer from the liquid-phase refrigerant flowing through the first refrigerant flow path 4 to the outer circumferential surface of the inner tube 3 was obtained for the cases of the inside diameter D of the inner tube 3 being 12 mm and 18 mm by computer simulation calculation which was conducted while the liquid flow path width W was varied under the following conditions: the elongated-projection thickness T2 of the elongated projections 13 of the inner tube 3 was set to 0.5 mm; the pitch P2 of the elongated projections 13 as measured at roots of the elongated projections 13 was set to 3 mm; the inside diameter D of the inner tube 3 was set to 12 mm and 18 mm; the wall thickness of the inner tube 3 was set to 1.2 mm; and the temperature and pressure of the liquid-phase refrigerant as measured at the inlet of the first refrigerant flow path 4 were set to 40° C. and 1.38 MPaG, respectively. FIG. 11 shows the results of the computer simulation calculation. It has been derived from the results shown in FIG. 11 that the a liquid flow path width W of 0.4 mm to 1.2 mm inclusive is preferred. As is apparent from FIG. 11, when the liquid flow path width W is less than 0.4 mm, pressure loss increases sharply. Also, when the liquid flow path width W is in excess of 1.2 mm, the overall heat transfer coefficient in heat transfer from the liquid-phase refrigerant flowing through the first refrigerant flow path 4 to the outer circumferential surface of the inner tube 3 drops. In FIG. 11, pressure loss and the overall heat transfer coefficient in heat transfer from the liquid-phase refrigerant flowing through the first refrigerant flow path 4 to the outer circumferential surface of the inner tube 3 are expressed in percentage with those of the case where the inside diameter D of the inner tube 3 is 12 mm, and the liquid flow path width W is 0.4 mm, being each taken as 100%.

FIG. 12 shows a modified inner tube of the double-wall-tube heat exchanger of Embodiment 1.

An inner tube 30 shown in FIG. 12 is twisted about its axis, so that the interior fins 12 and the elongated projections 13 are spiral.

Embodiment 2

The present embodiment is shown in FIGS. 13 to 17.

FIGS. 13 to 17 show the configurations of essential portions of a double-wall-tube heat exchanger according to Embodiment 2 of the present invention.

In the case of the double-wall-tube heat exchanger 31 of Embodiment 2, the outer tube 2 has a plurality of elongated projections 32 formed integrally with the inner circumferential surface of the outer tube 2. The elongated projections 32 project radially inward, extend in the longitudinal direction, and are arranged at equal circumferential intervals. Also, the outer tube 2 does not have expanded tube portions at opposite end portions thereof.

The inner tube 3 has diameter-reduced tube portions 33 located slightly longitudinally inward of the opposite ends thereof. A refrigerant inlet (not shown) is formed in the wall of the outer tube 2 at a position corresponding to one diameter-reduced tube portion (not shown). The refrigerant outlet 8 is formed in the wall of the outer tube 2 at a position corresponding to the other diameter-reduced tube portion 33. An end portion of a liquid-phase refrigerant inflow pipe (not shown) made of aluminum is inserted into the refrigerant inlet and is brazed to the outer tube 2. An end portion of the liquid-phase refrigerant outflow pipe 11 made of aluminum is inserted into the refrigerant outlet 8 and is brazed to the outer tube 2. Elongated projections are not formed on the outer circumferential surface of the inner tube 3.

Portions of the outer tube 2 which are located longitudinally outward of the diameter-reduced tube portions 33 of the inner tube 3 are subjected to roller working which is performed from the radial outside toward the radial inside along the entire circumference, thereby forming the diameter-reduced portions 14. The diameter-reduced portions 14 are brazed to portions of the inner tube 3 near the opposite ends thereof. The diameter-reduced portions 14 are formed after the inner tube 3 is placed within the outer tube 2. In the course of formation of the diameter-reduced portions 14, associated portions of the elongated projections 32 of the outer tube 2 are crushed and bite into the outer circumferential surfaces of the inner tube 3 (see FIGS. 16 and 17). This reduces the clearance between the outer circumferential surface of the inner tube 3 and a portion of the inner circumferential surface of the outer tube 2 where the elongated projections 32 are not formed, to such an extent as to be filled with a brazing material. In this condition, the diameter-reduced portions 14 of the outer tube 2 are brazed to the inner tube 3. Also, the clearance between the outer circumferential surface of the inner tube 3 and a portion of the inner circumferential surface of the diameter-reduced portion 14 of the outer tube 2 where the elongated projections 32 are not formed is filled with the brazing material 17 (see FIG. 17).

In contrast to the double-wall-tube heat exchanger 1 of Embodiment 1, when the expanded pipe portions 15 a and 16 a of the vapor-phase refrigerant inflow and outflow pipes 15 and 16 are to be fitted onto and brazed to respective opposite end portions of the inner tube 3, the double-wall-tube heat exchanger 31 of Embodiment 2 does not require an operation of cutting off relevant portions of the elongated projections 32 from the outer tube 2 and an operation of pressing the expanded pipe portions 15 a and 16 a of the vapor-phase refrigerant inflow and outflow pipes 15 and 16 from the radial outside. Even in the double-wall-tube heat exchanger 31 of Embodiment 2, preferably, brazing between the inner tube 3 and each of the expanded pipe portions 15 a and 16 a of the vapor-phase refrigerant inflow and outflow pipes 15 and 16 is carried out simultaneously with brazing between the outer tube 2 and the inner tube 3 while an appropriate gap is maintained between the opposite ends of the outer tube 2 and the corresponding ends of the expanded pipe portions 15 a and 16 a of the vapor-phase refrigerant inflow and outflow pipes 15 and 16.

Other configurational features are similar to those of the double-wall-tube heat exchanger 1 of Embodiment 1. The double-wall-tube heat exchanger 31 of Embodiment 2 is incorporated into the refrigeration system shown in FIG. 7 in a manner similar to that of the double-wall-tube heat exchanger 1 of Embodiment 1.

When the subcooled liquid-phase refrigerant flows through the liquid-phase refrigerant inflow pipe and flows into the first refrigerant flow path 4 of the double-wall-tube heat exchanger 31, by the effect of the unillustrated one diameter-reduced tube portion of the inner tube 3, the liquid-phase refrigerant dividedly flows into all the channels formed between the adjacent elongated projections 32 in the first refrigerant flow path 4. Having passed through all the channels formed between the adjacent elongated projections 32 in the first refrigerant flow path 4 of the double-wall-tube heat exchanger 31, flows of the liquid-phase refrigerant join together in the other diameter-reduced tube portion 33. The resultant liquid-phase refrigerant flows to the expansion valve 26 through the liquid-phase refrigerant outflow pipe 11.

The double-wall-tube heat exchanger 31 of Embodiment 2 may also use the inner tube 3 which is twisted about its axis.

FIG. 18 shows a modified inner tube of the double-wall-tube heat exchanger of Embodiment 2.

Opposite end portions of an inner tube 35 shown in FIG. 18 are extended, so that the vapor-phase refrigerant inflow pipe and the vapor-phase refrigerant outflow pipe are not used. Piping extending from the evaporator 25 is connected to an end portion of the inner tube 35 which is located on a side toward the refrigerant outlet 8. Piping extending to the compressor 20 is connected to the other end portion of the inner tube 35.

In Embodiment 1 described above, the elongated projections 13 are formed on the outer circumferential surface of the inner tube 3. In Embodiment 2 described above, the elongated projections 32 are formed on the inner circumferential surface of the outer tube 2. However, an embodiment may be such that the elongated projections 13 and the elongated projections 32 are formed on the outer circumferential surface of the inner tube 3 and the inner circumferential surface of the outer tube 2, respectively. In this case, the elongated projections 13 of the inner tube and the elongated projections 32 of the outer tube 2 are arranged in a circumferentially staggered manner. 

1. A double-wall-tube heat exchanger comprising an outer tube, and an inner tube disposed in and spaced apart from the outer tube, a clearance between the outer tube and the inner tube and the interior of the inner tube serving as respective refrigerant flow paths, the inner tube having a plurality of interior fins formed on an inner circumferential surface thereof, the interior fins projecting radially inward, extending in a longitudinal direction of the inner tube, and being arranged at circumferential intervals, the inner tube having a plurality of elongated projections formed on an outer circumferential surface thereof, the elongated projections projecting radially outward, extending in the longitudinal direction, and being arranged at circumferential intervals, and a fin height of the interior fins being greater than a projecting height of the elongated projections.
 2. A double-wall-tube heat exchanger according to claim 1, wherein a radial clearance between an inner circumferential surface of the outer tube and a portion of the outer circumferential surface of the inner tube where the elongated projections are not formed is 0.4 mm to 1.2 mm inclusive.
 3. A double-wall-tube heat exchanger according to claim 1, wherein a clearance between projecting ends of the elongated projections of the inner tube and the inner circumferential surface of the outer tube is 0 mm to 0.5 mm inclusive.
 4. A double-wall-tube heat exchanger comprising an outer tube, and an inner tube disposed in and spaced apart from the outer tube, a clearance between the outer tube and the inner tube and the interior of the inner tube serving as respective refrigerant flow paths, the inner tube having a plurality of interior fins formed on an inner circumferential surface thereof, the interior fins projecting radially inward, extending in a longitudinal direction of the inner tube, and being arranged at circumferential intervals, and the outer tube having a plurality of elongated projections formed on an inner circumferential surface thereof, the elongated projections projecting radially inward, extending in a longitudinal direction of the outer tube, and being arranged at circumferential intervals.
 5. A double-wall-tube heat exchanger according to claim 4, wherein a radial clearance between a portion of the inner circumferential surface of the outer tube where the elongated projections are not formed and an outer circumferential surface of the inner tube is 0.4 mm to 1.2 mm inclusive.
 6. A double-wall-tube heat exchanger according to claim 4, wherein a clearance between projecting ends of the elongated projections of the outer tube and the outer circumferential surface of the inner tube is 0 mm to 0.5 mm inclusive.
 7. A double-wall-tube heat exchanger according to claim 1, wherein the interior fins of the inner tube have a fin thickness of 0.2 mm to 2.0 mm inclusive.
 8. A double-wall-tube heat exchanger according to claim 1, wherein the interior fins have a fin height of 1.0 mm to 3.0 mm inclusive.
 9. A double-wall-tube heat exchanger according to claim 1, wherein the inner tube has an inside diameter of 12 mm or greater.
 10. A double-wall-tube heat exchanger according to claim 4, wherein the interior fins of the inner tube have a fin thickness of 0.2 mm to 2.0 mm inclusive.
 11. A double-wall-tube heat exchanger according to claim 4, wherein the interior fins have a fin height of 1.0 mm to 3.0 mm inclusive.
 12. A double-wall-tube heat exchanger according to claim 4, wherein the inner tube has an inside diameter of 12 mm or greater. 